Working fluid circuit for active suspension control system of vehicle

ABSTRACT

A working fluid circuit for an active suspension control system for an automotive vehicle comprises a pressure source unit which provides pressurized working fluid to actuators, associated with wheels respectively, which are operable to suppress attitude change of a vehicle body, pressure control valves associated with the actuators respectively, and check valves located in a by-pass from an output port to a supply port of the pressure control valve, the supply port communicating with the pressure source through a supply line to which the working fluid is supplied from the pressure source, the output port communicating with the actuator through an output line for outputting the working fluid adjusted at a preselected controlled level to the actuator. Each of the check valves is responsive to pressure of the working fluid in the output line which is greater than that in the supply line for feeding back the working fluid in the output line to the supply line for providing a part of pressure supplied to said pressure control valve.

This application is a continuation of application Ser. No. 07/787,868filed Nov. 5, 1991 now abandoned.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates generally to an active suspension controlsystem for an automotive vehicle. More specifically, the inventionrelates to an improved working fluid circuit for an active suspensioncontrol system which is operable to provide control pressure effectivelywith reduced consumption of energy.

2. Description of the Background Art

Japanese Patent First Publication No. 62-295714 discloses an activesuspension control system for an automotive vehicle. This systemincludes working cylinders each disposed between a vehicle body and awheel, pressure control valves responsive to command signals to controlworking fluid pressure to be supplied to the working cylindersrespectively, means for detecting or projecting a lateral accelerationacting on the vehicle body, and a controller responsive to a signalindicative of a lateral acceleration value to calculate command signalsto the pressure control valves. The controller is operable to multiplythe lateral acceleration value by control gains to provide the commandsignals which serve to suppress rolling motion of the vehicle.

In such a conventional system, working fluid at high pressure must bealways supplied to a supply port of a pressure control valve such as anelectromagnetic proportional pressure reducing valve for creatingpressure in a working fluid cylinder against inertial force of rollingmotion of a vehicle. Therefore, a great supply flow rate is necessarywhich is provided by a pressure source including a hydraulic pump and areservoir tank for supplying high energy to the working fluid cylinders,resulting in great consumption energy. It will be noted that a vehicleincorporating a conventional active suspension control system consumesmore fuel than a vehicle without an active suspension control system dueto the additional energy consumed for anti-rolling motion control.Accordingly, an improved active suspension control system which conductsthe same level of control as that of the conventional system, but withreduced energy consumption, has been sought.

SUMMARY OF THE INVENTION

It is therefore a principal object of the present invention to avoid thedisadvantages of the prior art.

It is another object of the invention to provide a working fluid circuitfor an active suspension control system which is effectively operablewith reduced necessary energy to be supplied by a pressure source,without degrading active suspension operation.

According to one aspect of the present invention, there is provided aworking fluid circuit for an active suspension control system for avehicle which comprises a pressure source providing pressurized workingfluid to actuators associated with wheels respectively, which actuatorsare operable to suppress attitude change of a vehicle body, pressurecontrol valves controlling pressure of the working fluid supplied by thepressure source to the actuators respectively, each having a supply portcommunicating with the pressure source through a supply line throughwhich the working fluid is supplied from the pressure source to thepressure control valve and an output port communicating with theactuator through an output line through which the working fluid,adjusted at a preselected controlled level, is output to the actuator,and valve means associated with the pressure control valvesrespectively, each being responsive to pressure of the working fluid inthe output line which is greater than that in the supply line tofeedback the working fluid in the output line to the supply line forproviding pressure as a part of supply pressure to said pressure controlvalves.

In the preferred mode, the valve means includes by-pass lines each ofwhich communicates between the output line and the supply line connectedto the pressure control valve and check valves disposed in the by-passlines respectively.

Additionally, the valve means may further include throttles eacharranged in the by-pass line in series with the check valve forrestricting working fluid of a pressure above a preselected pressurelevel from flowing to the supply line connected to the pressure controlvalve.

With the above arrangement, when, for example, the inside wheels travelover a rough shoulder or road protrusion, during turning, vibration isinput and a working pressure for the inside wheels (i.e., pressure ofthe output port of the pressure control valve or controlled pressure) iselevated above a supply pressure. The valve means is opened to allow theworking fluid to flow from the output line to the supply line throughthe by-pass line. This raises pressure of the working fluid supplied tothe pressure control valve. It will be thus noted that vibrationalenergy caused by the protrusions of the road surface is converted intofluid energy to add to the supply pressure to the supply port of thepressure control valve for recovering reduction in supply pressure foruse in anti-rolling control necessary due to energy consumption of theoutside wheels. Therefore, the supply pressure provided by the pressuresource may be decreased by the recovered pressure.

BRIEF DESCRIPTION OF THE DRAWINGS

In the drawings:

FIG. 1 is an illustration which shows an active suspension controlsystem according to the present invention.

FIG. 2 is a sectional view which shows a structure of a pressure controlvalve for use in an active suspension control system.

FIG. 3 is a graph which shows the relationship between output pressureof a pressure control valve and a command signal or current appliedthereto.

FIG. 4 is graph which shows timing charts representing relations betweensupply pressures and control pressures for right and left wheelsrespectively in a conventional active suspension control system.

FIG. 5 is a graph which shows timing charts representing relationsbetween supply pressures and control pressures for right and left wheelsrespectively in the present active suspension control system.

FIG. 6 is a block diagram which shows an alternate embodiment of anactive suspension control system of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENT

Referring now to the drawings, particularly to FIG. 1, an activesuspension control system 10 for use in an automotive vehicle accordingto the present invention is shown. This system 10 generally includes ahydraulic pump 12 and reservoir tank 14 as a pressure source, amulti-valve assembly 16, a front pressure control valve assembly 18F forfront wheels, a rear pressure control valve assembly 18R for rearwheels, actuators or hydraulic cylinders 20FL to 20RR, and a controlunit 22 which provides control command signals for suppressing vehicleattitude change.

The pressure source provides pressurized working fluid such as hydraulicfluid to the multi-valve assembly 16. The multi-valve assembly 16 thensupplies the hydraulic fluid to the front and rear pressure controlvalve assemblies 18F and 18R to control pressure to be supplied to thehydraulic cylinders 20FL to 20RR for active suspension control.

The hydraulic pump 12 includes a plunger type pump for example which isdriven by rotational output from a vehicle engine. A inlet port of thepump 12 is communicated with the reservoir tank through a pressure line23 and an outlet port thereof is communicated with a supply line 24which connects an oil filter 26 and a check valve 28 serially. The checkvalve 28 restricts working fluid flow from backing-up to the hydraulicpump 12. The supply line 24 branches from the multi-valve assembly 16.The branched lines are communicated with the front and rear pressurecontrol valve assemblies 18F and 18R respectively. In addition, a pumpaccumulator 29 is provided in the supply line 24.

Return lines 30 extend from the pressure control assemblies 18F and 18Rand merge at the multi-valve assembly 16 for communicating with anoperational check valve 32 and a throttle 34 serially. The merged returnline 30 is then connected to the reservoir tank 14 through an oil cooler36.

The operational check valve 32 is designed as a pilot-operated checkvalve to which supply pressure output from the check valve 28 is fed aspilot pressure Pp. This check valve 32 is operable to maintain an openposition when the pilot pressure Pp is greater than a preselected level,such as neutral pressure, and to assume a closed position in response toa pilot pressure Pp less than the preselected level. With thisarrangement, closing of the operational check valve 32 due to enginestop causes line pressure in a hydraulic circuit upstream from theoperational check valve 32 to be held at the preselected level.

The multi-valve assembly 16 includes a relief valve 38 which has beenset to a preselected relief pressure and is disposed between the checkvalve 28 and the operational check valve 38 to connect the supply line24 downstream from the check valve and the return line 30.

The supply line 24 introduced into the front pressure control valveassembly 18F branches to be communicated with supply ports of front-leftand front-right pressure control valves 40FL and 40FR as describedhereinafter in detail. Return lines connected to return ports of thepressure control valves 40FL and 40FR, are merged to define the returnline 30 which then extends to the multi-valve assembly 16. Additionally,the rear pressure control valve assembly 18R includes pressure controlvalves 40RL and 40RR and a line network thereof is comparable with thefront pressure control valve assembly 18F, therefore, its arrangementwill not described in detail here.

Referring to FIG. 2, one example of the pressure control valves 40FL to40RR is shown. This pressure control valve generally includes acylindrical valve housing 43 incorporating a valve body and aproportional solenoid 44 integrally provided in the valve housing.

The valve housing 43 includes a cylindrical valve bore 43A at its centerportion in which a main spool 45 is slidably disposed, and, a pilotvalve bore 43B which is formed at a longitudinal end portion of thevalve bore 43A coaxially therewith. A poppet is slidably disposed in thepilot valve bore 43B as a pilot valve. In both ends of the main spool45, a feed-back chamber FR and a pilot chamber PR are definedrespectively. Springs 47A and 47B are arranged in the chambers FR and PRrespectively which serve to center the spool 45 in the valve bore 43A. Athrottle 43Aa is provided by a support disposing the upper spring 47Bfor restricting a flow rate of the working fluid passing through thepilot chamber PR and the pilot valve bore 43B.

The valve housing 43 further includes a supply port 43s, a return port43r, and an outlet port 43o which are communicated with the valve bore43A and disposed so as to face toward lands 45a and 45b of the mainspool 45 and a pressure chamber 45c respectively. A valve seat 43Bawhich defines an opening having a preselected diameter is provided, inthe pilot valve bore 43B, opposite a tip portion of the poppet 46.

The supply port 43s is communicated with the pilot valve bore 43B via asupply side passage 48. The return port 43r is also communicated withthe pilot valve bore 43B through a return side passage 49. With thesepassages, the working fluid from the supply port 43s is partlycirculated to the return port 43r through the passage 48, the valve seat43Ba, the passage 49. During this circulation, the return side passageis communicated with the pilot valve bore 43B at both ends of the poppet46 and also communicated with the inside of the proportional solenoid44.

The output port 43o is communicated with the feed-back chamber FRthrough a feed-back passage 50. Between the feed-back chamber FR and aportion below the poppet 46, as viewed from the drawing, in the pilotvalve bore 43B, a by-pass passage 52 is formed which bypasses the spoolvalve bore 43A. Disposed in the by-pass passage 52 is a check valve 54which serves to allow the working fluid to flow from the spool valvepassage 43A to the feed-back chamber FR.

The proportional solenoid 44 includes an axially slidable plunger 58 andan exciting coil 59 for moving the plunger 58 in an axial direction ofthe valve 44. The exciting coil 59 is responsive to a command signal orcurrent i to be excited for moving the plunger 58, urging the poppet 46to contact with the valve seat 43B. The degree of displacement of thepoppet toward the valve seat 43B determines a flow rate of the workingfluid through the valve seat 43B for adjusting pressure in the pilotchamber PR to a preselected level. Characters a-e in the drawingindicate throttles.

It is noted that when the pressures in the feed-back chamber FR and thepilot chamber PR are balanced under the condition that thrust from theproportional solenoid 44 acts on the poppet 46, the spool 45 takes anoverlapping position (shown in FIG. 2) for blocking communicationbetween the output port 43o and the supply port 43s. The pressure in thepilot chamber PR (i.e., the pilot pressure) is modified according to themagnitude of the command current i. The spool 45 then vibrates slightlyuntil the pilot pressure and the pressure in the feed-back chamber FRare balanced with each other. This controls an output pressure Pc(control pressure) from the output port 43o proportionally to themagnitude of the command current i as shown in FIG. 3. In FIG. 3, P₂designates a maximum line pressure to be supplied.

In a case where a vibration of a low frequency in a spring resonanceregion (for example, around 1 Hz) is transmitted from a road surface,the spool 45 can be slightly displaced so that the working fluidreciprocates between the hydraulic pressure source and a load side toabsorb vibration in pressure within a preselected range.

The output ports 43o of the pressure control valves 40FL-40RR arerespectively connected with output lines 62, and the output ports 43oare communicated with the cylinder chambers of the hydraulic cylinders20FL-20RR by the output lines 62 as shown in FIG. 1.

Further, the front and rear pressure control valve assemblies 18F and18R of this embodiment include by-pass lines 64 which connect betweenthe output lines 62 and the supply lines 24 by-passing the pressurecontrol valves 40FL-40RR respectively. Junctions of the by-pass lines 64and the output lines 62 are arranged close to the hydraulic cylinderswith respect to the throttles c of the output ports 43o.

Each by-pass line 64 includes a check valve 66 which serves as a passageopen and close mechanism which is responsive to a preselected pressurelevel in the output line 62 to feed back the working fluid in the outputline 26 to the supply line 24. In FIG. 1, front and rear accumulators 68are provided in the supply lines 24 extending to the front and rearpressure control valve assemblies 18F and 18R.

The hydraulic cylinders 20FL-20RR serving as actuators are provided withsingle acting cylinders each disposed between a wheel and a vehiclebody. This cylinder includes a cylinder chamber L into which workingfluid pressure is adjusted by the pressure control valve through whichits pressure is supplied. The cylinder chamber L is communicated with athrottle 70 and an accumulator 72 which serve to absorb variation inpressure in a sprung resonance range. Additionally, coil springs (notshown) are interposed between the wheels and the vehicle bodyrespectively for supporting a static load of the vehicle body.

The control unit 22 includes generally a lateral acceleration sensor 74,a vehicle height sensor 75, and a controller 76. The controller 76incorporates a micro-computer and is responsive to a signal Y_(G)representing a lateral acceleration acting on the vehicle body, inputfrom the lateral acceleration sensor 74 to multiply it by a preselectedanti-roll control gain to calculate command currents i to be output tothe pressure control valves 40FL-40RR associated with the wheelsrespectively. The controller 76 is further responsive to a signal hindicating a vehicle height monitored by the vehicle height sensor 75 toprovide command currents so that the vehicle body is maintained at alevel orientation within a target vehicle height range. These commandcurrents are interposed in the command currents i which are, in turn,output to the pressure control valves 40FL-40RR as active suspensioncontrol signals.

In operation, assuming that a vehicle is running straight at a constantspeed with a standard load, the lateral acceleration sensor 74 providesa signal Y_(G) of zero which indicates no acceleration acting on thevehicle body. The controller 76 is then responsive to signals h from thevehicle height sensor 75 to provide the command signals i (e.g., a valuei_(N) corresponding to the neutral pressure P_(N) shown in FIG. 3) tothe pressure control valves 40FL-40RR respectively for maintaining avehicle height level within the target height range. The pressurecontrol valves 40FL-40RR are responsive to the command signals i tocontrol pressures to be supplied to the hydraulic cylinders 20FL-20RR,establishing cylinder pressures equal to P_(N). As a result, a cylinderstroke of the hydraulic cylinder is set according to the cylinderpressure P_(N) to maintain the vehicle body at a flat level orientationwithin the target height range.

When the vehicle is turned to the left from a straight running status,the vehicle body tends to roll and the lateral acceleration sensor 74detects lateral acceleration acting on the vehicle body to provide asignal Y_(G) indicating the magnitude of the lateral acceleration. Thecontroller 76 then increases command currents i to be supplied to thepressure control valves 40FR and 40RR associated with the right wheelswhich are the outside wheels, in accordance with a value Y_(G) of thelateral acceleration, while it decreases command currents i to besupplied to the pressure control valves 40FL and 40RL. Thus, thecylinder pressures of the hydraulic cylinders 18FR and 18RR for theoutside wheels are elevated from the present neutral pressure P_(N) forexample, while the cylinder pressures of the hydraulic cylinders 18FLand 18RL for the inside wheels are reduced from the neutral pressureP_(N), causing force in the hydraulic cylinders 18FR and 18RR to becreated against lowering motion due to an inertial force of the vehiclebody. Cylinder pressures of the hydraulic cylinders 18FL and 18RL forthe inside wheels are reduced to provide pressures which do not promoterolling motion. It will be appreciated that anti-rolling control iseffected in the vehicle body to restrict the rolling motion formaintaining vehicle attitude on the horizontal plane.

Conventionally, assuming that the vehicle with a heavy load is runningstraight on an even road at a constant speed, due to the increase in aload, the vehicle height level is lowered, thus, the controller 76effects vehicle height control to increase a control pressure P_(C) (forexample, 80 kgf/cm² =P_(N) ' as shown in FIG. 4), when vehicle attitudeor anti-rolling motion control is not carried out, toward a level whichis higher than the above-described control pressure P_(C) (for example,50 kgf/cm² =P_(N)) in the case of a standard load, resulting in apressure difference relative to the supply pressure P_(S) (for example,100 kgf/cm² as shown in FIG. 4) being reduced so that the degree ofpressure control to outer wheels may become small or inadequate.

In order to clarify the object of the present invention, pressurevariation, when the vehicle is heavily loaded, in a conventional systemwithout the bypass lines 64 with the check valves 66 for each of thepressure control valves 40FL-40RR, will be described hereinbelowreferring to FIG. 4.

If a vehicle is running straight with a heavy load in the same manner asdescribed above. Although a high control pressure P_(C) (=80 kgf/cm²) isutilized as a neutral pressure P_(N) ' when the vehicle is runningstraight, the controller 76 adjusts the vehicle body to be maintained ata level of a target vehicle height. When the vehicle is temporarilyturned to the left while running straight, by the above describedrolling control, control pressures P_(C) (i.e., cylinder pressures) ofthe hydraulic cylinders 40FR and 40RR for the outside wheels rise asshown in FIG. 4, while control pressures P_(C) of the hydrauliccylinders 40FL and 40RL for the inside wheels are made to be reduced.

However, referring to FIG. 4, due to supply of working fluid to thehydraulic cylinders for the outside wheels, the supply pressure P_(S)begins to be reduced after a time t₁. Thus, after a time t₂ when thesupply pressure P_(S) coincides with the control pressure P_(C), acontrol pressure does not correspond to a command pressure (which isshown by a broken line in FIG. 4). The control pressure P_(C) is loweredin line with the supply pressure P_(S).

On the other hand, a command for decreasing the control pressure P_(C)for the left or inside wheels is output after the time t₁ as shown bybroken line. Between the time t1 and time t₂, control pressure P_(C) isvaried according to the decrease of the supply pressure P_(S).

During the rolling motion control beginning at the time t1, if a vehicletravels over a protrusion in a road surface and vibration of relativelylarge amplitude caused thereby is input only to the inside wheels (theleft wheels) at the time t₂ or a time close to t₂, the working fluid inthe cylinder chamber L is compressed to cause an internal pressure ofthe cylinder chamber L to increase over a command pressure as shown inthe drawing.

It will be noted that when large vibration is inputted to the hydrauliccylinder, caused by traveling over a large protrusion or traveling at ahigh speed, the control pressure P_(C) for the left wheels becomeshigher than the supply pressure P_(S) between a time t₃ and t₄ as shownby a hatched region in FIG. 4. In other words, vibrational energytransmitted from a road surface is converted into hydraulic energy,causing the control pressure P_(C) to elevate above the supply pressureP_(S) (P_(C) >P_(S)). Accordingly, it will be appreciated that hydraulicenergy necessary to be generated by the suspension system of a vehicleitself can be reduced by utilizing vibrational energy transmitted from aroad surface effectively.

Referring to FIG. 5, pressure variation according to the system of theinvention under the same conditions as described above (a heavy load, aleft turn, and traveling over a protrusion in a road surface by leftwheels) is shown.

When the left wheel (inside wheel) travels over a protrusion of a roadsurface at a time t₁, vibrational energy is input to the hydrauliccylinder to increase control pressure P_(C) for the left wheel. At atime t₂ when the control pressure P_(C) is greater than supply pressureP_(S) at a time t₃ as shown in FIG. 5, the check valve 66, for example,is opened, which by-passes the pressure control valve 40FL for thefront-left wheel, allowing the working fluid in the cylinder chamber Land in the output line communicated with the cylinder chamber L to bereturned to the supply port 43s through the check valve 66. This resultsin the supply pressure P_(S) starting to elevate.

Upon the supply pressure P_(S) being elevated, it is utilized forsuppressing rolling motion at the right wheels. Therefore, the supplypressure P_(S) varies according to a curve, as shown in FIG. 5, whichextends through the middle of the excess portion of the curve indicatingthe control pressure P_(C) over the supply pressure P_(S).

It will be appreciated that hydraulic energy converted from vibrationalinput to the left wheels which corresponds to the hatched region A istransferred for control of the right wheels so that control pressureP_(C) is elevated from a pressure level in a conventional system byenergy corresponding to the hatched region C (=to the hatched region A).Accordingly, an insufficiency of control pressure P_(C) relative to atarget pressure becomes less than that of the supply pressure of aconventional system, which is shown by one-dotted-chain-line, enhancinganti-rolling motion control. In contrast, control pressure P_(C) for theleft wheels is lowered from a conventional control pressure by a degreedefined by the hatched region A (=the hatched region C).

As mentioned above, the present invention is directed to a simplestructure which incorporates the bypass line 64 and the check valve 66into a conventional active suspension control system. With thisstructure, hydraulic energy corresponding to hatched regions "A+B" inFIG. 5 can be utilized effectively. Actual control pressures P_(C) foreither the right or left wheels are provided which are close todesirable pressure levels aimed by the controller 76. This results inimproved favorable control characteristics.

It can be said from a different viewpoint that the same controlperformance as that of the conventional active suspension control systemcan be secured even when supply flow (rated flow) provided by thehydraulic pump 12 is somewhat reduced. Therefore, hydraulic energy to besupplied the pump to the hydraulic circuit can be reduced by a degree ofincrease in hydraulic energy generated by road surface vibrationalinput, thus improving fuel consumption. It will be appreciated that thesuspension of the present invention can be easily applied to theconventional systems with low manufacturing cost.

As shown in FIGS. 4 and 5, control pressure P_(C) for the left wheel isoscillated by rebounding motion caused after the wheel has passed over aroad protrusion.

In the system of the invention, when vibrational energy input isextremely high and supply pressure P_(S) exceeds a preselected setpressure level of the relief valve 38 due to energy conversion from thevibrational energy input, working fluid or hydraulic fluid, an amount ofwhich corresponds to the excess of the supply pressure P_(S), isreturned to the reservoir tank 14 through the relief valve 38 sincelocation of the relief valve is close to the load side, therebyrestricting the maximum value of supply pressure P_(S) generated by theenergy conversion, preventing the pressure control valves 40FL to 40RRfrom being damaged.

Although in this embodiment, a typical example has been described inwhich only the inside wheels run over protrusions during anti-rollingcontrol, it can been seen easily that when the outside wheels run overprotrusions simultaneously with the inside wheels, an insufficiency ofsupply energy is covered in the same manner as already mentioned so thatvibrational energy from a road surface is converted effectively intohydraulic energy.

Referring to FIG. 6, an alternative embodiment is shown. This embodimentincludes throttles or orifices 78 each disposed in a by-pass line 64upstream from a check valve 66 in a serial relation thereto. Otherarrangements are the same as those of the above described firstembodiment. As to elements common to the embodiment shown in FIG. 1,like numbers refer to like parts and description thereof will beomitted.

According to this embodiment, the same working effect as that describedbefore can be obtained. Further, when excessive high energy is inputtedfrom a road surface, a sharp variation in supply pressure caused by theenergy conversion is restricted by the orifice 78 to a preselectedpressure level for preventing the working fluid of pressure above apreselected level from flowing to the supply line for advantageouslyestablishing further stable operation of a pressure control valve.

The above embodiments include the by-pass lines 64 and the check valves66 which are provided separately from the pressure control valves 40FLto 40RR respectively. However, these arrangements may be integrallyprovided within the pressure control valves.

Additionally, the flow passage opening and closing mechanism may beprovided with a pressure sensor and an electromagnetic directionalcontrol valve in place of the check valve 66. With these arrangements,the pressure sensor monitors supply and cylinder pressures to providesignals indicative thereof to the controller 76. The controller comparesthe magnitude of these pressure levels to control the directionalcontrol valve according to the result of comparison for opening andclosing the by-pass line 64.

Further, the above described active suspension control system mayutilize a gas, such as air, as a working fluid in place of hydraulicfluid.

As already mentioned, a by-pass passage is provided which communicates asupply port with an output port of a pressure control valve under thecondition that a spool is by-passed by the by-pass passage, wherein theby-pass passage is provided with a passage opening and closing mechanismsuch as a check valve which allows hydraulic fluid to flow from theoutput port to the supply port when pressure (control pressure) of theoutput port is greater than pressure (supply pressure) of the supplyport. Accordingly, for example, when supply pressure is greatly loweredand high vibration or high amplitude input is given from a road surfaceunder the condition that the vehicle is turned, the control pressureexceeds the supply pressure, and hydraulic fluid passes through theby-pass passage so that it can be sent to a supply line. This increasesthe supply pressure and rolling control effect can be improved, so thatvibrational energy given from a road surface can be converted into fluidenergy effectively to add pressure to the supply port of the pressurecontrol valve for covering reduction in supply pressure for use inanti-rolling control caused by energy consumption for outside wheels.The supply pressure provided by a hydraulic pressure source may bedecreased by the covered pressure. As a result, the capacity of thehydraulic pressure source may be reduced as compared with a conventionalsystem, so that energy consumption of the vehicle can be reduced withimprovement in active suspension control.

What is claimed is:
 1. A working fluid circuit for an active suspensioncontrol system for a wheeled vehicle at least three wheels, comprising:aplurality of actuators cooperating with respective wheels of the vehiclefor suppressing an attitude change of the vehicle body; a pressuresource for providing a pressurized working fluid to said actuators; aplurality of pressure control valves for respectively controlling apressure of the working fluid supplied by said pressure source to eachof said actuators, each pressure control valve having a supply port andan output port; a plurality of supply lines connecting said pressuresource and respective supply ports of said pressure control valves forsupplying pressurized working fluid from said pressure source tocorresponding pressure control valves, said supply lines being in fluidunrestricted communication with each other; a plurality of output linesconnecting between respective output ports of said pressure controlvalves and corresponding actuators for adjusting the working fluid ineach of said actuators at preselected controlled levels; and valve meanscomprising respective single check valves connecting each of said supplylines and said output lines, each of said single check valves beingprovided in parallel to a respective one of said pressure controlvalves, said single check valves each acting when a pressure of theworking fluid in the output line connected to said respective pressurecontrol valve exceeds a pressure in the supply line connected to saidrespective pressure control valve to feed back working fluid from theoutput line connected to said respective pressure control valve to thesupply line connected to said respective pressure control valve tothereby supplement the supply pressure of the working fluid provided viasaid unrestrictedly communicating supply lines to all of said pressurecontrol valves.
 2. A circuit as set forth in claim 1, wherein said valvemeans includes by-pass lines each of which communicates between theoutput line and the supply line connected to the pressure control valveand check valves disposed in the by-pass lines respectively.
 3. Acircuit as set forth in claim 2, wherein said valve means furtherincludes throttles each arranged in the by-pass line in series with thecheck valve for restricting working fluid above a preselected pressurelevel from flowing to the supply line connected to the pressure controlvalve.
 4. A circuit as set forth in claim 1, wherein each of saidpressure control valves is connected to said pressure source through areturn line for returning the working fluid to said pressure source, acheck valve being disposed in the supply line which restricts workingfluid flow from returning to said pressure source and a relief valvewhich communicates between the supply line downstream from the checkvalve and the return line, the relief valve being responsive to pressureof the working fluid in the supply line which is greater than apreselected pressure level to allow the working fluid to flow to saidpressure source through the return line.
 5. A circuit as set forth inclaim 1, wherein each of said pressure control valves includes a spool,pilot pressure being supplied to an end of the spool, the controlledpressure output from the output port being fed back to the other end ofthe spool, said pressure control valve being responsive to the pilotpressure to adjust a pressure level of the working fluid supplied fromsaid pressure source.